Pulsation attenuation in systems with multiple compression elements

ABSTRACT

In compression systems having multiple sources of pulsation induced noise, provisions are made to reduce the noise by selectively varying the length of adjacent refrigerant lines leading to a common refrigerant manifold. Each of the refrigerant lines is associated with a particular noise source. The feature can be incorporated into compressors operating in parallel or in a single multi-rotor compressor having multiple suction or/and discharge ports. It is equally applicable to both discharge and suction ports.

FIELD OF THE INVENTION

This invention relates generally to compressors and, more particularly,to a method and apparatus for reducing pulsation induced noise generatedby compression elements operating in parallel.

BACKGROUND OF THE INVENTION

Typical air conditioning, heat pump and refrigeration systems include atleast one compressor for circulating a refrigerant throughout the systemwhich includes at least a condenser (or a gas cooler), an expansiondevice and an evaporator. In such vapor compression systems, thecompressor is a major source of pulsation induced noise. Duringoperation, the compressor generates refrigerant pulsations propagatingthroughout the refrigerant system and contributes about fifty percent tothe overall system sound level. These pulsations transmitted through therefrigerant system can result in damage and distraction of the systemcomponents and can be disruptive to occupants of a conditioned space.

The types of compressors that are used in heating, ventilation, airconditioning and refrigeration (HVAC&R) applications include, forexample, reciprocating compressors, rotary compressors, scrollcompressors and screw compressors. Screw compressor installations areparticularly recognized to generate pulsation induced noise.

It is common in HVAC&R applications to have multiple compressorinstallations with compressors operating in parallel (or in tandem) orhave a plurality of compression elements that may be driven, forexample, by a common shaft. In these applications, pulsation inducednoise can be particularly disruptive and bothersome, since the amplitudeof pulsations may be additive from these multiple sources.

The HVAC&R equipment manufacturers have made efforts to reduce pulsationinduced noise levels generated by refrigerant systems implementingvarious techniques such as sound attenuating materials, active noisecontrol and the like. Many of these methods involve additional cost andcomplexity or have unproven effectiveness and reliability record. Also,for the occupant comfort and safety and protection of the environment,various governmental regulations and legislative measures imposinglimits on the sound levels emanating from HVAC&R systems have beenimplemented. Therefore, there is a need for efficient and inexpensivemethods for sound attenuation and pulsation reduction for multiplecompressors or compression elements operating in parallel.

DISCLOSURE OF THE INVENTION

In an HVAC&R system having a plurality of pulsation induced noisesources associated with compressor discharge/suction ports, connected toa common manifold, such that the lengths of the connecting ducts areselectively varied in such a way that the pressure pulsations in therespective ducts cancel each other at the point where they reach thecommon manifold.

By one aspect of the invention, a plurality of compressors operating inparallel have discharge ducts of varying lengths which are connected toa common outlet manifold.

By another aspect of the invention, a compressor, such as a screwcompressor or scroll compressor, including at least a pair of dischargeports has discharge ducts of differing lengths that interconnect to acommon discharge manifold.

By yet another aspect of the invention, a duality of discharge/inletducts vary in length in accordance with the equation L_(Dif)=n*(c/f),where c is the speed of sound in the refrigerant, f is the dominantpulsation frequency and n is an integer number.

In the drawings as hereinafter described, a preferred embodiment isdepicted; however, various other modifications and alternateconstructions can be made thereto without departing from the spirit andscope of the invention.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic illustration of a vapor compression system havingmultiple compressors with the present invention incorporated therein.

FIG. 2 is a graphic illustration of the two sound waves emanating fromtwo sound sources and cancelling each other.

FIG. 3 is a graphic illustration of three sound waves emanating fromthree compressors or compression elements in accordance with the presentinvention.

FIG. 4 is a graphic illustration of a tri-rotor screw compressor withdual discharge ports to which the present invention is applicable.

DETAILED DESCRIPTION OF THE INVENTION

Shown in FIG. 1 is a schematic illustration of a basic vapor compressionsystem 11 which includes, in a serial flow relationship, a condenser 12,an expansion device 13, an evaporator 14 and a compression system 16. Inthe cooling or refrigeration mode of operation, refrigerant flowsserially from the compression system 16 to the condenser 12, theexpansion device 13, the evaporator 14 and back to the compressionsystem 16 which includes dual compressors 17 and 18 operating in tandem.In a heat pump mode of operation, the functions of the evaporator 14 andthe condenser 12 are reversed. Obviously, by incorporation ofappropriate flow control devices such as a four-way valve, the vaporcompression system 11 can be made capable to run in both cooling andheating modes of operation.

As will be seen, the compression system 16 includes two compressors 17and 18 arranged in parallel with a common suction manifold 19 and acommon discharge manifold 21. The compressor 17 is connected to thecommon suction manifold 19 by a refrigerant suction line 22 and to thecommon discharge manifold 21 by a refrigerant discharge line 23.Similarly, the compressor 18 is fluidly connected to the suctionmanifold 19 by a refrigerant suction line 24 and to the common dischargemanifold 21 by the refrigerant discharge line 26. Although only twocompressors 17 and 18 operating in tandem are shown in FIG. 1, a largernumber of compressors can be included into the compression system 16.

Sound pressure waves tend to emanate from the compressors 17 and 18,most significantly from their respective discharge ports 27 and 28, butalso from their suction ports 29 and 31. The present invention isimplemented to minimize the generated noise effect of these pressurepulsations.

Recognizing that the pressure pulsations are typically transmitted byplane wave propagation, the respective lengths of the refrigerantdischarge line 23 and the refrigerant discharge line 26 are varied insuch a manner as to cause the respective pressure pulsations to canceleach other out at the point where they enter the common dischargemanifold 21. The same is true for the lengths of the refrigerant suctionline 22 and the refrigerant suction line 24 where the pressurepulsations travel toward the common suction manifold 19. In order forthis wave cancellation to occur, the lengths of the refrigerantdischarge lines 23 and 26 and/or the lengths of the refrigerant suctionlines 22 and 24 are selected such that the difference in length isL_(Dif)=n*(c/f), where c is the speed of sound in the particularrefrigerant, f is the dominant pulsation frequency and n is an integernumber.

As an example, a typical speed of sound in a refrigerant would vary from150 m/s to 250 m/s and a typical dominant pulsation frequency would bein the range of 30 Hz to 600 Hz Thus, L_(Dif) would vary from 0.25 m to8.3 m (with n=1).

Shown in FIG. 2 is a graphic illustration of the sound waves amplitudeas a function of time at the point in each refrigerant linecommunicating with each compressor or compression element 17 and 18within the compression system 16, where the refrigerant lines locatedeither on the suction or discharge side, merge together into the commonsuction or discharge manifold respectively. Each of these sound wavesrepresents the pulsation at this location as emanating from each of thetwo compressors. The sound wave shown as “x” is the wave emanating fromthe first compressor 17, and the sound wave shown as “y” is the waveemanating from the compressor 18. In this case, each of those tworefrigerant lines are chosen to be of a different length, such that eachsound wave arrives at the junction point to be 180 degrees out of phasewith the other wave. Due to of the difference in the lengths of therefrigerant lines, as described hereinabove, the two sound waves canceleach other out.

As mentioned above, it should be recognized that the present inventionis applicable not only to dual compressors but to the compressionsystems including multiple compressor in general. For example, in athree compressor configuration, the discharge refrigerant lines (and/orthe suction refrigerant lines) are of three different lengths for thepurpose discussed hereinabove in accordance to the following formulaL_(Dif)=(2/3)*n*(c/f).

In a three compressor configurations, the respective sound waves thatresult from the differing duct lengths thus correspond by the wave formsA, B and C as shown in FIG. 3. Again, at the point where the sound wavesarrive to the common discharge manifold 21, or common suction manifold19, they cancel each other out.

These embodiments would be especially applicable to the compressionsystems with multiple compressors whose rotating speed of operation issynchronized. In a more general sense, for multiple synchronouspulsation sources connected to a common manifold to cancel each other,the length of each corresponding refrigerant line can be offset byL_(Dif)=(2/k)*n*(c/f), where k is the number of pulsation sources orcompressor ports.

Referring now to FIG. 4, a basic vapor compression system 60 includes ina serial flow relationship a tri-rotor compressor 32, a condenser 33, anexpansion device 34, and an evaporator 36. An economizer circuit, withadditional vapor injection and unloading capability, may be included butis not shown.

The tri-rotor compressor 32 includes a centrally located drive rotor 37and a pair of driven rotors 38 and 40 on either side thereof. The driverotor 37 may be driven by an electric motor (not shown), and in turndrives the driven rotors 38 and 40. Compression chambers are definedbetween the screw flutes on the drive rotor 37 and the respective driverotors 38 and 40. Refrigerant, which is compressed in the compressionchambers between the rotors, is discharged through discharge ports 39and 41, to respective refrigerant discharge line 42 and 43 and finallyto a discharge manifold 44 prior to passing to the condenser 33.

The present invention is incorporated in such a tri-rotor system byselectively varying the lengths of the two discharge ducts 42 and 43 ina manner described herein above.

On the suction side of the tri-rotor compressor, refrigerant flows fromthe evaporator 36 to an inlet manifold 46, to the refrigerant inletlines 47 and 48, and then to the respective inlet ports 49 and 51. Therespective lengths of the refrigerant inlet lines 47 and 48 can beselectively chosen so as to cancel out their respective pressurepulsations emanating from the inlet ports 49 and 51 in a manner asdescribed hereinabove.

Once again, this invention is applicable to the compressors includingmultiple compression elements, where the lengths of the refrigerantlines leading to suction or/and discharge port for each compressionelement is established in accordance to the formula disclosedhereinabove for multiple compressors. Also, although the invention isdescribed with respect to a tri-rotor screw compressor, other compressortypes, such as root compressors or scroll compressors connected to thesame shaft, are within the scope and can equally benefit from theinvention.

Furthermore, although the invention is disclosed with respect to thesound level attenuation, other benefits, such as for instance vibrationlevel reduction or sound quality improvement, can be also obtained.These and other benefits of the invention could be recognized by aperson ordinarily skilled in the art.

It should be understood that this invention applies to a broad range ofrefrigerant systems, including container refrigeration units,truck-trailer systems, residential cooling units and heat pumps, rooftop installations and the like.

While the present invention has been particularly shown and describedwith reference to a preferred mode as illustrated in the drawings, itwill be understood by one skilled in the art that various changes indetail may be effected therein without departing from the spirit andscope of the invention as defined by the claims.

1. Apparatus for at least one of reducing pulsation induced noise,reducing vibration level or improving sound quality in an HVAC&R systemhaving a plurality of compressor ports, comprising: a plurality ofrefrigerant lines with each refrigerant line fluidly interconnecting oneof said plurality of compressor ports to a common refrigerant manifold;wherein said plurality of refrigerant lines are of selectively differinglengths so as to result in pressure pulsations cancellation. 2.Apparatus as set forth in claim 1 wherein said ports are compressordischarge ports and said refrigerant lines are discharge lines. 3.Apparatus as set forth in claim 1 wherein said ports are compressorsuction ports and said refrigerant lines are suction lines.
 4. Apparatusas set forth in claim 1 wherein said plurality of compressor portscomprise two ports and further wherein the difference in lengths of saidrefrigerant lines is determined by the equation L_(Dif)=n*(c/f), whereinc is the speed of sound of refrigerant, f is the dominant pulsationfrequency, and n is an integer number.
 5. Apparatus as set forth inclaim 1 wherein said plurality of ports comprise a plurality of portswithin a plurality of compressors connected in parallel.
 6. Apparatus asset forth in claim 1 wherein said plurality of ports comprises aplurality of ports in a single compressor.
 7. Apparatus as set forth inclaim 6 wherein said single compressor is one of a screw compressor, ascroll compressor, a rotary compressor or a root compressor. 8.Apparatus as set forth in claim 1 wherein said HVAC & R system includesat least a compressor, an evaporator and an expansion device. 9.Apparatus as set forth in claim 1 wherein the differences in lengths ofsaid refrigerant lines are determined by the equationL_(Dif)=(2/1c)*n*(c/f), wherein c is the speed of sound of refrigerant,f is the dominant pulsation frequency, k is the number of pulsationsources, and n is an integer number.
 10. A method for at least one ofreducing pulsation induced noise, reducing vibration level or improvingsound quality in an HVAC&R system having a plurality of compressor portsthat are susceptible of producing pressure pulsations, comprising thesteps of: providing a plurality of refrigerant lines with eachrefrigerant line fluidly interconnecting one of said ports to a commonrefrigerant manifold; wherein said plurality of refrigerant lines are ofselectively different lengths so as to result in their respectivepressure pulsations canceling each other out at the point of theirreaching said common refrigerant manifold.
 11. A method as set forth inclaim 10 wherein said ports are compressor discharge ports and saidrefrigerant lines are discharge lines.
 12. A method as set forth inclaim 10 wherein said ports are compressor suction ports and saidrefrigerant lines are suction lines.
 13. A method as set forth in claim10 wherein said plurality of compressor ports comprise two ports andfurther wherein the difference in length of said refrigerant lines isdetermined by the equation L_(Dif)=n*(c/f), wherein c is the speed ofsound of refrigerant, f is the dominant pulsation frequency, and n is aninteger number.
 14. A method as set forth in claim 10 wherein saidplurality of ports comprise a plurality of ports within a plurality ofcompressors connected in parallel.
 15. A method as set forth in claim 10wherein said plurality of ports comprises a plurality of ports in asingle compressor
 16. A method as set forth in claim 10 wherein thedifferences in lengths of said refrigerant lines are determined by theequation L_(Dif)=(2/k)*n*(c/f), wherein c is the speed of sound ofrefrigerant, f is the dominant pulsation frequency, k is the number ofpulsation sources, and n is an integer number.